Internally cooled high compression lean-burning internal combustion engine

ABSTRACT

An internally cooled internal combustion piston engine and method of operating a piston engine is provided, with the combination of liquid water injection, higher compression ratios than conventional engines, and leaner air fuel mixtures than conventional engines. The effective compression ratio of the engines herein is greater than 13:1. The engines may employ gasoline or natural gas and use spark ignition, or the engines may employ a diesel-type fuel and use compression ignition. The liquid water injection provides internal cooling, reducing or eliminating the heat rejection to the radiator, reduces engine knock, and reduces NOx emissions. The method of engine operation using internal cooling with liquid water injection, high compression ratio and lean air fuel mixture allow for more complete and efficient combustion and therefore better thermal efficiency as compared to conventional engines.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a continuation application of co-pending U.S. patentapplication Ser. No. 14/949,523 filed Nov. 23, 2015, which is acontinuation application of U.S. patent application Ser. No. 14/598,935filed Jan. 16, 2015, now U.S. Pat. No. 9,194,339 issued Nov. 24, 2015,which is a continuation application of U.S. patent application Ser. No.13/444,533 filed Apr. 11, 2012, now U.S. Pat. No. 8,935,996 issued Jan.20, 2015, which has been filed as U.S. Reissue patent application Ser.No. 15/410,356 filed Jan. 19, 2017, which application claims priorityunder 35 U.S.C. §119(e) of U.S. Provisional Patent Application No.61/474,240, filed Apr. 11, 2011, the disclosure of which is herebyincorporated by reference in its entirety.

BACKGROUND

The present disclosure pertains to the field of internal combustionengines, including engines for motor vehicles, railways, ships,aircraft, or electrical power generation.

This disclosure pertains to internal combustion engines that operate farmore efficiently than conventional engines. The principles set forthherein can be used in both spark-ignition (SI) engines typicallyoperating on gasoline (petrol), ethanol or natural gas, or oncompression-ignition engines, which typically are diesel engines.

The engine literature describes a number of factors that affect engineefficiency. These can be divided into theoretical limits based on thesecond law of thermodynamics, namely the temperature differential(gradient) that determines efficiency in the Carnot cycle, andcompression ratio which is the most pertinent variable in Otto cycleefficiency. Other factors are important, including mechanical factorssuch as friction and chemical factors such as fuel properties. Fuelproperties depend on the chemical makeup of the fuel, the stoichiometry,vaporization of liquid fuels, and other factors, including thecombustion temperature, ignition energy and ignition delay, flamepropagation velocity, and completeness of combustion.

Internal combustion engines are heat engines, whose behavior can bedescribed in the ideal limit by the laws of thermodynamics. The work andthermal energy of any heat driven process can be described by the firstlaw of thermodynamics as:

Q _(in) =W _(out) +Q _(out)

where Q_(in) is thermal energy put into the engine, and mechanicalenergy, or work is W_(out). A cyclic heat engine, even in the ideallimit, cannot completely convert the net heat input into work output, sosome of the input heat energy has to be dissipated into the environmentas waste heat Q_(out) The thermal efficiency of a cyclic heat engine isdefined as:

${\eta_{th} \equiv \frac{W_{out}}{Q_{i\; n}}} = {1 - \frac{Q_{out}}{Q_{i\; n}}}$

where η_(th) is a dimensionless efficiency factor. This is a performancemeasure of a device that uses thermal energy, such as an internalcombustion engine.

The theoretical maximum efficiency of any heat engine is given by theCarnot theorem, which posits that the theoretical maximum efficiency ofany heat engine depends on the difference between hot and coldtemperature reservoirs in an ideal thermodynamically reversible engine.This maximum efficiency in a Carnot engine is defined to be:

$\eta_{th} \leq {1 - \frac{T_{C}}{T_{H}}}$

where T_(C) is the absolute temperature of the cold reservoir, and T_(H)is the absolute temperature of the hot reservoir. Therefore, efficiencyin a Carnot engine is a factor of the temperature gradient between thehot and cold reservoirs.

The Otto cycle is another ideal thermodynamic cycle that relates engineefficiency of internal combustion spark-ignition engines to compressionratio. The geometry of Otto cycle employs two adiabatic and two constantvolume processes. Otto cycle efficiency, which assumes perfect gas lawbehavior, can be expressed as:

$\eta_{th} = {1 - \frac{1}{r^{\gamma - 1}}}$

where r is the volume compression ratio, and γ=Cp/Cv, the specific heatratio, of heat capacity at constant pressure (C_(P)) to heat capacity atconstant volume (C_(V)). A similar formula for diesel engines relatescompression ratio (and combustion expansion ratio) to efficiency indiesel (compression ignition) engines. The specific heat ratio is alsoknown as the “isentropic expansion factor.” The specific heat ratio ofthe air-fuel mixture γ varies with temperature and the heat capacity ofthe fuel vapor, but is generally close to the air value of 1.4. Whenusing this standard value, the cycle is called an “air-standard cycle.”Because γ is always greater than 1, engine efficiency in the Otto cycleis directly related to compression ratio. Therefore, high compressionratio engines will operate more efficiently than a lower compressionratio engine, all other factors being equal.

Temperature control in engines is also an important factor affectingengine efficiency. The Carnot cycle suggests that the higher thetemperature after the ignition at top dead center (TDC) of the piston inthe cylinder (i.e., the highest temperature in the engine), the largerthe temperature differential will be, which leads to greater efficiency.However, real world inefficiencies include the lack of complete mixingof the fuel with air, the rate of combustion, and the air/fuel ratiorequired for effective ignition. Most engines operate at a nearstoichiometric air/fuel ratio. Combustion under these conditions createsexcess heat that is not converted into mechanical work. This excess heatmust be rejected using a radiator or through the exhaust. The highcombustion temperatures created also create undesirable NOx emissions.

Temperature control in modern engines is usually accomplished by acooling jacket surrounding the engine, transporting heat to a heatexchanger (radiator) that rejects excess heat to the environment andmaintains the engine within operating temperature limits. The use of aconventional radiator in such a fashion is termed herein as externalcooling. Most modern internal combustion engines are liquid (or water)cooled (externally cooled) using either water or some other liquidcoolant, which circulates through the engine and runs through the heatexchanger. Alternatively, some engines are characterized as “aircooled,” typically because they lack a radiator. Instead, most aircooled engines have additional fins integral with the engine block orcylinders to convect and radiate heat away from the engine.

Even in the most efficient liquid or air cooled conventional engines,the requirement to shed heat through the cooling system significantlydecreases engine efficiency. Approximately 40% of engine heat isdissipated in the radiator or cooling fins, which is lost energy, somepart of which could still theoretically be useable as mechanical energy.Thus, reducing this heat loss, and converting excess heat to usefulmechanical energy, is an important unmet need in engine design.Conventional automobiles are only about 20% efficient at converting theenergy in gasoline to mechanical energy. The remaining 80% or so of theenergy in the fuel is lost to the environment through the cooling systemand heat exchanger (radiator) and as exhaust heat. Thus, if the heatloss through the radiator (or otherwise dissipated to the environment)could be substantially reduced, engine efficiency could be substantiallyimproved.

The compression ratio in engines which use fuels like gasoline ornatural gas is limited by the need to control engine knock, which iscaused by pre(auto)-ignition of the fuel prior to the desired ignitionfrom firing of the spark plug. During pre-ignition, fuel ignites duringthe compression stroke in an uncontrolled fashion due to the hightemperatures generated in the cylinder during compression. Suchpre-ignition wastes energy and could lead to engine damage ifuncontrolled. To avoid engine knock, conventional spark ignition enginesare generally limited to an effective compression ratio of about 10:1,with up to 12:1 possible with more expensive high octane fuel.

An additional factor affecting engine performance is the air (oxygen) tofuel ratio. Stoichiometric air provides one mole of molecular oxygen permole of carbon and 0.5 mole of molecular oxygen per mole of hydrogen inthe fuel. The amount of air for true stoichiometric oxygen is dependenton the exact chemical makeup of the fuel, but is approximately 14.7:1weight/weight (w/w) for gasoline and diesel engines (i.e., 1 gram offuel to 14.7 grams of air). Engines are typically run rich during coldstart and high load operation, but when run rich, there will benon-combusted fuel and thus wasted energy and additional air pollution.Engines normally run most efficiently at about a stoichiometric mixture,but there are theoretical bases for efficient engine operation underlean conditions, of greater than stoichiometric oxygen.

SUMMARY

In one aspect, there is provided a system and method for operating aspark or compression-ignition engine at elevated compression ratioscompared to conventional engines, using lean air fuel ratios and liquidwater injection to control the temperature inside the cylinder duringboth the compression and power strokes. The higher compression allowsfor higher thermal efficiency in accordance with the Otto or compressionignition (diesel) ideal engine cycles, and also allows for reliableignition of leaner fuel mixtures. The liquid water injection reduces thework during compression by reduction in pressure, controls knock andprovides temperature control. The liquid water injection also reducesthe need for external cooling leading to less heat loss to the radiatorand thus higher efficiency. The combination of liquid water injectionand other heat management features, including using very lean air/fuelmixtures, can eliminate totally or substantially reduce the need for aradiator (i.e., a smaller radiator can be employed) than conventionalengines, and have substantially lower heat losses to the environment.Engines described herein therefore produce much higher thermalefficiency compared to conventional engines, as well as lower emissions.

According to one embodiment, there is provided an internal combustionengine for use with a hydrocarbon fuel, with at least one cylinder and areciprocating piston therein, at least one air intake valve providingair into the at least one cylinder, at least one exhaust valve, and afuel handling system with a fuel injector providing fuel into the atleast one cylinder, comprising a water injector coupled to a liquidwater source for injecting liquid water (by direct injection) into thecylinder at any time from about 180° to about 30° before TDC during acompression stroke of the piston, wherein the amount of liquid waterinjected is greater than the amount of water that is present at thesaturation point of water vapor in the ambient air in the cylinder.Alternatively, the engine may be provided with an air intake manifold influid communication with water or fuel injectors or both, such that thewater or fuel or both are port injected into the intake manifold ratherthan directly into the cylinder. In this port injection embodiment, thewater injector will be controlled to inject liquid water at a somewhatearlier time in the cycle, typically from about 300° to about 180°before TDC when the intake valve is still open. The engines have a ratioof air to fuel provided to the at least one cylinder that is greaterthan stoichiometric.

The engines have an effective compression ratio greater than 13:1. In anembodiment, the engine has an effective compression ratio greater than15:1. In an embodiment, the compression ratio can be as high as 20:1, orhigher. For example, in spark-ignition engines typically operating ongasoline (petrol) ethanol or natural gas, the compression ratio rangesfrom about 13:1 to about 25:1 and in another embodiment, from about 13:1to about 20:1. In other embodiments, it is, about 16:1 or about 17:1 orabout 18:1 or about 19:1 or about 20:1 or about 21:1 or about 22:1 orabout 23:1 or about 24:1 or about 25:1. In diesel fuel engines, in anembodiment the compression ratio may be lower, for example, from about12:1 or about 13:1, including all of the ranges and values describedherein above, but, in addition, in other embodiments it is higher, e.g.,it may range up to about 35:1 or higher, for example, about 25:1, orabout 26:1 or about 27:1 or about 28:1 or about 29:1 or about 30:1 orabout 31:1, or about 32:1, or about 33:1, or about 34:1 or about 35:1.

In an embodiment, when liquid water is directly injected into thecylinder, the injection is timed to occur at from about 180° to about30° before TDC during a compression stroke of the piston. The internalcombustion engine of the present invention includes, in one embodiment,a water injector for direct injection into the cylinder, and in anotherembodiment water injector for port injection into the cylinder, and inanother embodiment two water injectors, one directly into the cylinderand the other through a port. It is to be understood that direct waterinjection can occur at any time during the cycle from about 180° toabout 30° before TDC during a compression stroke of the piston. Moreoverthe water injection may be at the same or at a different position in thecompression stroke of the piston, from cycle to cycle of the compressionstroke. For example in one stroke, it may be at position of about 600before TDC, and in another cycle at about 90° before TDC, the timing andthe amount being controlled, as described herein below. In anembodiment, the water injection may range from about 90° to about 60°before TDC, for example, when the water is directly injected into thecylinder.

The above-description is also applicable when the liquid water is portinjected. In an embodiment when port injected; however, the liquid watercan be port injected at a value outside of 30° to 180° before TDC, suchas, for example, at about 300° to about 180° before TDC.

Further in an embodiment, an amount of liquid water injected in anengine cycle ranges at about 1.05 to about 10 times the amount of watervapor carried by air saturated with water vapor at ambient temperatureof 25° C. at the engine intake.

Alternately, in another embodiment, an amount of liquid water injectedin an engine cycle is about 20% to about 800% w/w of the amount of fuelbeing injected in the engine cycle. In one embodiment, controls areimplemented such that at higher compression ratios, the greater is this% difference. The controls implemented maintain compression endtemperature at specified value. As in-cylinder temperature is notmeasured, the control may be implemented through ambient pressure,temperature, humidity and in-cylinder pressure relative to engine loadand engine RPM (rotations per minute).

According to a further aspect, there is provided a method of operatingan internal combustion engine for use with a hydrocarbon fuel, theengine having at least one cylinder and a reciprocating piston therein,at least one air intake valve providing air into the at least onecylinder, at least one exhaust valve, and a fuel handling system with afuel injector providing fuel into the at least one cylinder. The methodcomprises: injecting liquid water into the cylinder at any time fromabout 180° to about 30° before TDC of the piston during compression,wherein the amount of liquid water injected is greater than the amountof water that is present at the saturation point of water vapor in theambient air in the cylinder, wherein a ratio of air to fuel provided tothe at least one cylinder is greater than stoichiometric; and, whereinthe engine operates at an effective compression ratio of greater than13:1.

Further to this embodiment, the method includes injecting an amount ofliquid water in an engine cycle of about 1.05 to about 10 times anamount of water vapor carried by air saturated with water vapor atambient temperature of 25° C. at the engine intake.

Further to this embodiment, the method includes injecting an amount ofliquid water in an engine cycle of about 20% to about 800% w/w of theamount of fuel in the engine cycle.

As described herein in an embodiment the liquid water is injected intothe cylinder any time from about 180° to about 300 before TDC. Inanother embodiment, water is injected any time from about 45° to about120° before TDC, and in another embodiment, from about 600 to about 90°before TDC. It is understood that any value or range from about 180° toabout 30° before TDC is contemplated, for example, 180°, 179°, 178°,177°, 176°, 175°, 174°, 173°, 172°, 171°, 170°, 169°, 168°, 167°, 166°,165°, 164°, 163°, 162°, 161°, 160°, 159°, 158°, 157°, 156°, 155°, 154°,153°, 152°, 151°, 150°, 149°, 148°, 147°, 146°, 145°, 144°, 143°, 142°,141°, 140°, 139°, 138°, 137°, 136°, 135°, 134°, 133°, 132°, 131°, 130°,129°, 128°, 127°, 126°, 125°, 124°, 123°, 122°, 121°, 120°, 119°, 118°,117°, 116°, 115°, 114°, 113°, 112°, 111°, 110°, 109°, 108°, 107°, 106°,105°, 104°, 103°, 102°, 101°, 100°, 99°, 98°, 97°, 96°, 95°, 94°, 93°,92°, 91°, 90°, 89°, 88°, 87°, 86°, 85°, 84°, 83°, 82°, 81°, 80°, 79°,78°, 77°, 76°, 75°, 740, 73°, 72°, 71°, 70°, 69°, 68°, 67°, 66°, 65°,64°, 63°, 62°, 61°, 60°, 59°, 58°, 570, 56°, 55°, 54°, 53°, 52°, 51°,50°, 49°, 48°, 47°, 46°, 45°, 44°, 43°, 42°, 41°, 40°, 39°, 38°, 37°,36°, 35°, 34°, 33°, 32°, 31°, 30°, before TDC.

If the water is port injected, the injection may occur from about 3000to about 1800 before TDC. It is understood that any value or range fromabout 180° to about 300° before TDC contemplated, for example, 300°,299°, 298°, 297, 296°, 295°, 294°, 293°, 292°, 291°, 290°, 289°, 288°,287°, 286°, 285°, 284°, 283°, 282°, 281°, 280°, 279°, 278°, 277°, 276°,275°, 274°, 273°, 272°, 271°, 270°, 269°, 268°, 267°, 266°, 265°, 264°,263°, 262°, 261°, 260°, 259°, 258°, 257, 256°, 255°, 254°, 253°, 252°,251°, 250°, 249°, 248°, 247°, 246°, 245°, 244°, 243°, 242°, 241°, 240°,239°, 238°, 237, 236°, 235°, 234°, 233°, 232°, 231°, 230°, 229°, 228°,227°, 226°, 225°, 224°, 223°, 222°, 221°, 220°, 219°, 218°, 217, 216°,215°, 214°, 213°, 212°, 211°, 210°, 209°, 208°, 207, 206°, 205°, 204°,203°, 202°, 201°, 200°, 199°, 198°, 197°, 196°, 195°, 194°, 193°, 192°,191°, 190°, 189°, 188°, 187°, 186°, 185°, 184°, 183°, 182°, 181°, or180° before TDC.

As defined hereinabove, in an embodiment, the amount of liquid waterinjected ranges from about 1.05 times to about 10 times the amount ofwater vapor carried by air saturated with water vapor at ambienttemperature of 25° C. at the engine intake. Thus for example, in variousembodiments, the amount of water injected may be at any values in therange described hereinabove or may range from about 1.05 to about 10times the amount of water vapor carried by air saturated by water vaporat ambient temperature at the engine intake such as, for example, 1.25,1.50, 1.75, 2.00, 2.25, 2.50, 2.75, 3.00, 3.25, 3.50, 3.75, 4.00, 4.25,4.50, 4.75, 5.00, 5.25, 5.50, 5.75, 6.00, 6.25, 6.50, 6.75, 7.00, 7.25,7.50, 7.75, 8.00, 8.25, 8.50, 8.75, 9.00, 9.25, 9.50, 9.75, or 10.00.times the amount of water vapor carried by air saturated by water vaporat ambient temperature of 25° C. at the engine intake.

In alternative embodiment, the amount of water injected ranges fromabout 20% to about 800% w/w of the fuel. Any range or value from about20% to about 800° w/w fuel can be utilized, e.g., 25%, 30%, 35%, 40%,45%, 50%, 55%, 60%, 65%, 70%, 75%, 80%, 85%, 90%, 95%, 100%, 105%, 110%,115%, 120%, 125%, 130%, 135%, 140%, 145%, 150%, 155%, 160%, 165%, 170%,175%, 180%, 185%, 190%, 195%, 200%, 205%, 210/0%, 215%, 220%, 225%,230%, 235%, 240%, 245%, 250%, 255%, 260%, 265%, 270%, 275%, 280%, 285%,290%, 295%, 300%, 305%, 310%, 315%, 320%, 325%, 330%, 335%, 340%, 345%,350%, 355%, 360%, 365%, 370%, 375%, 380%, 385%, 390%, 395%, 400%, 405%,410%, 415%, 420%, 425%, 430%, 435%, 440%, 445%, 450%, 455%, 460%, 465%,470%, 475%, 480%, 485%, 490%, 495%, 500%, 505%, 510%, 515%, 520%, 525%,530%, 535%, 540%, 545%, 550%, 555%, 560%, 565%, 570%, 575%, 580%, 585%,590%, 595%, 600%, 605%, 610%, 615%, 620%, 625%, 630%, 635%, 640%, 645%,650%, 655%, 660%, 665%, 670%, 675%, 680%, 685%, 690%, 695%, 700%, 705%,710%, 715%, 720%, 725%, 730%, 735%, 740%, 745%, 750%, 755%, 760%, 765%,770%, 775%, 780%, 785%, 790%, 795%, or 800%.

In another embodiment, the amount of water injected range from about 40%to about 400% (w/w) of the amount of fuel being injected in the enginecylinder.

In another embodiment, the amount of water injected range from about 50%to about 300% (w/w) of the amount of fuel being injected in the enginecylinder.

In still another embodiment, the amount ranges from about 60% to about200% (w/w) of the amount of fuel being injected in the engine cylinder.

The loss of dissipated heat in the internal combustion engine isminimized by utilizing water injection in which the amount of liquidwater injected is greater that in conventional combustion engines.

BRIEF DESCRIPTION OF THE DRAWINGS

The objects, features and advantages will become apparent to oneordinary skill in the art, in view of the following detailed descriptiontaken in combination with the attached drawings, in which

FIG. 1A show a cutaway view of the configuration and liquid water spraypattern of the liquid water injector in an example cylinder in oneembodiment, and FIG. 1B is an underside view of the cylinder head orchamber taken along line A-A of FIG. 1A;

FIG. 2 depicts a matrix 500 describing the alternative combinations ofinternal combustion engine configurations for providing liquid waterinjection and heat management features as described herein;

FIG. 3 depicts an example configuration of an internal combustion engine50 of a first embodiment having liquid water injection features asdescribed herein;

FIG. 4 depicts an example configuration of an internal combustion engine150 of a second alternative embodiment having liquid water injectionfeatures as described herein;

FIG. 5 depicts an example configuration of an internal combustion engine250 of a third alternative embodiment having liquid water injectionfeatures as described herein;

FIG. 6 depicts an example configuration of an internal combustion engine350 of a fourth alternative embodiment having liquid water injectionfeatures as described herein;

FIG. 7 depicts an example configuration of an internal combustion engine450 of a fourth alternative embodiment having liquid water injectionfeatures as described herein;

FIG. 8 illustrates a sensor control system 100 implemented in thevarious alternative internal combustion engines described herein in oneembodiment;

FIG. 9 illustrates a method implemented by an engine control unit fordetermining an amount of water to be injected per cycle in the variousalternative internal combustion engines described herein in oneembodiment;

FIG. 10 depicts a plot of engine efficiency at various engine loads,with water injection and supercharged air injection pressure; and

FIG. 11 shows engine efficiency at various engine loads and superchargeair injection, with and without water injection.

DETAILED DESCRIPTION

This disclosure provides two-stroke or four-stroke, or higher strokecombustion engines with at least one cylinder employing direct liquidwater cylinder injection and/or port injection features to regulate thetemperature of the combustion process, as the combustion is in progress.The engine may thus include a spark plug, or a glow plug, a plasmaigniter, or a laser igniter providing for spark ignition, plasmaignition, pilot ignition, laser ignition, free radical ignition or sparkassist compression ignition and operate with gasoline (petrol), alcoholor combination thereof or natural gas as fuel. Alternatively, the enginemay employ compression ignition, such as a diesel (kerosene) poweredengine with or without additional assistance of spark, plasma or laser.

FIG. 2 shows a matrix 500 of various embodiments of an internalcombustion engine including direct liquid water injection features andother heat management techniques described herein. For example, each ofsixteen (16) combinations 502 show various engine configurations withmarks (e.g., an “X”) indicating: a combustion engine 505 e.g., sparkignition engines (or alternatively, by omission of an “X” mark,indicates a compression type engine); presence of a turbocharger element508 (indicating use of a turbocharger device); implementation of adirect liquid water injection in the cylinder 512 or liquid waterinjection at the inlet port 522 (e.g., by suction); implementation ofdirect fuel injection in the cylinder 515 or fuel injection at the inletport 525. Column 510 shows exemplary embodiments described herein. Forexample, as will be described in greater detail, FIG. 3 shows an enginecombination of a first embodiment indicated as 50, FIG. 4 shows anengine combination of an alternative embodiment indicated as 150, FIG. 5shows an engine combination of an alternative embodiment indicated as250, FIG. 6 shows an engine combination of an alternative embodimentindicated as 350, and FIG. 7 shows an engine combination of analternative embodiment indicated as 450. Like reference numbers in thevarious views indicate like elements.

FIGS. 3-7 illustrate some of the embodiments described in FIG. 2.However, contemplated within the scope of the present disclosure areengines comprised of direct water and direct fuel injection, alone or incombination with port water or port fuel injectors, in either aspark-ignited engine or diesel engine. By the term “direct injection” ismeant that the fuel or water being injected is injected directly intothe cylinder. In the case of direct fuel injection, the fuel is injectedwithout premixing with air. By the term “port injection” is meant thatthe fuel or water are injected into an intake manifold, where premixingwith air occurs prior to the air/vapor/water mix entering the cylinder.In some embodiments, the engine, whether spark-ignited or diesel, isturbocharged or supercharged. Also contemplated are engines comprised ofdirect water injection or direct fuel injection or both and/oralternatively port water or port fuel injection, or both, or anycombination thereof as long as the engine has at least one means forinjecting water and injecting fuel.

Referring to FIG. 3, the internal combustion engine 50 of a firstembodiment includes a fuel injection system 55 including a fuelreservoir 57 providing fuel to a fuel pump 53 which supplies fuel viafluid conduit or like transport means to a fuel injector device 56mounted to the cylinder head portion 20 to provide fuel under controlledconditions of temperature and pressure for combustion in the cylinder12. Fuel injector 56 injects fuel directly into the cylinder withoutpremixing with air. The fuel injection system can include a means forcontrolling the timing of fuel injection. Under timed computer control,the fuel can be charged into the cylinder.

In FIG. 3, the internal combustion engine 50 of a first embodimentincludes a liquid water injection system 65 including a liquid waterreservoir 67 providing water to a water pump 63 which supplies liquidwater via a fluid conduit or liquid transport device to a liquid waterinjector device 46 mounted to the cylinder head portion 20 to providedirect liquid water injection under controlled conditions of timing,pressure (e.g. variable or constant pressure) into the cylinder 12.Injector 46 as embodied in FIG. 3 injects water directly into thecylinder. The control system described below may control the injectionof liquid water into the cylinder 12 at one or more timed instances,every compression cycle. As shown in FIG. 3, an ignition coil 37controls firing of spark plug 47, mounted, in one embodiment, near thecenter of the cylinder head portion, situated between the liquid waterinjector 46 and the fuel injector device 56 mounted in the cylinder headportion 20.

In the embodiment of FIG. 3, in communication with cylinder head portion20 is an engine intake valve 21 timely actuated at each cycle to provideair from an intake manifold 25 for combustion with fuel within cylinderhead portion 20. Likewise, in communication with the cylinder headportion 20 is an exhaust valve 31 timely actuated at each cycle toenable exhaust gas products from combustion (carbon dioxide, air or anyother emissions) to exit the cylinder to an exhaust manifold 35, where,in one embodiment, is captured to perform further work for the engine,e.g., heated air.

In a further embodiment, shown in FIG. 3, a turbocharger sub-system 75is provided at air input, e.g., a manifold inlet 24, for receiving boththe input ambient air 11, and receiving from exhaust manifold portion 36hot exhaust gases 91 from the products of combustion to form aturbocharged compressed air mixture for combustion. In an alternativeembodiment, a supercharger may be used instead. In either case, theturbocharger or supercharger is controllable to adjust the amount of airforced into the cylinder or intake manifold.

As shown in the embodiment of engine 50 in FIG. 3, there is provided astructure and methodology of recapturing the heat byproduct of internalcombustion engines that can be used to further heat the air or fuel orliquid water at the cylinder. For example, the turbo-compressed airmixture 19 at the output of the turbocharger 75 is subject to heatregulation, e.g., heat removal via heat exchanging sub-system 70 thatincludes a heat exchange device 71 for recapturing a substantial portionof the waste heat produced by combustion and converting it into usefulenergy. In one aspect, heated exhaust gases 91 are re-circulated fromthe exhaust manifold 35 via manifold extension 36 for input to theturbocharger element 75 and input to heat exchanger 71 for use inpre-heating the liquid water to be injected and/or for use inpre-heating the air/fuel to be injected. As shown in FIG. 3, heat energyfrom the re-circulated exhaust air 91 is controllably added undercontrol of valve 74, e.g., via conduit 73 to the heat exchanger 71 tothereby regulate temperature of intake air 29 provided to the cylinderhead portion 20 for combustion. Further heated gas exhaust is removedvia heat exchanger 71, e.g., via conduit 72. The controllably removedheat from the exhaust gases 91 may be used to pre-heat the liquid waterin water reservoir 67.

Further, in the embodiment of the engine 50 of FIG. 3, there is a waterrecovery unit 66, i.e., a unit that extracts water from the exhaust gasfor example by cooling the exhaust down to ambient temperature byconventional means, e.g., condenser or by the by use of nanoporemembranes in which the water is condensed from the exhaust streamthrough capillary action, and the like. Thus, for example, in anembodiment, a condenser is provided that will capture water from anywater vapor by-product from the exhaust gas 91 which water may input viaa fluid conduit or coupling 68 to the water reservoir 67. The use of awater recovery device 66 as described herein also serves the purpose ofreducing the water storage requirements for the engines. This may beparticularly important in, for example, automotive applications, wherethe amount of water that would be required on board the vehicle could besubstantial absent means to capture and recycle water in the exhauststream.

A further embodiment is shown in FIG. 4 which depicts a non-turbochargedengine 150 of an alternative embodiment that implements direct liquidwater and direct fuel injection into the cylinder head portion 20 viarespective injectors 46, 56. In the embodiment depicted in FIG. 4,ambient air 11 is input to a heat exchanger device 71 via inlet 24 andhot exhaust gases 91 are circulated to the heat exchanger device 71under control of valve device 94. The heat from the hot exhaust gases isused to pre-heat the air 29 that is input to the cylinder forcombustion. Cooled gases 92 are re-circulated back to the output exhaustmanifold 35 via exhaust manifold portion 36 for engine output. A waterrecovery unit 66 as defined herein is provided to capture liquid waterfrom any water vapor present in the exhaust gas 91 which water may inputvia a fluid conduit or coupling 68 to the water reservoir 67.

In another embodiment, FIG. 5 depicts a turbocharged or superchargedengine 250 that implements port liquid water injecting and port fuelinjection via respective injectors 46, 56 into respective ports 38, 39formed at the intake manifold 25 associated with the cylinder 12. Thatis, respective liquid water supply system 65 provides water undercontrolled timing and (variable or constant) pressure conditions to portliquid water injector 46 at the port of intake manifold 25 near theintake valve 21 under control system operation. Likewise, fuel supplysystem 55 provides fuel under timed control and pressure conditions tothe fuel injector 56 at the port near the intake valve of intakemanifold 25. Otherwise, the embodiment of FIG. 5 is similar to engine 50depicted in FIG. 3. For example, engine 250 in FIG. 5, also provides forthe recapturing the heated gaseous product of combustion that can beused to further heat the air or fuel or liquid water at the cylinder.For example, the turbo-compressed air mixture 19 at the output of theturbocharger 75 is subject to heat regulation, e.g., heat removal viaheat exchanging sub-system 70 that includes a heat exchange device 71for capturing a substantial portion of the waste heat produced bycombustion and converting it into useful energy to control pre-heatingof the air/fuel and water. In one aspect, the heated exhaust gas 91 isre-circulated from the exhaust manifold 35 via manifold portion 36 forinput to the turbo charger element 75 and input to heat exchanger 71 foruse in pre-heating the liquid water to be injected and/or for use inpre-heating the air/fuel to be injected. As shown in FIG. 5, heat energyfrom the re-circulated exhaust air 91 is controllably added undercontrol of valve 74, e.g., via conduit 73 to the heat exchanger 71 tothereby regulate temperature of intake air 29 provided to the cylinderhead portion 20 for combustion. Further heated gas exhaust is removedvia heat exchanger 71, e.g., via conduit 72. The controllably removedheat from the exhaust gases 91 may be used to pre-heat the liquid waterin water reservoir 67. Further, in the embodiment of the engine 250 ofFIG. 5, a water recovery unit 66 is provided that will capture waterfrom any water vapor product from the exhaust gas 91 which water mayinput via a fluid conduit or coupling 68 to the water reservoir 67.

A further embodiment is shown in FIG. 6 which depicts an engine 350 thatimplements port liquid water injecting and port fuel injection viarespective injectors 46, 56 into respective ports 38, 39 formed at theintake manifold 25 associated with the cylinder 12. Otherwise, theembodiment of FIG. 6 is similar to engine 150 depicted in FIG. 4,wherein ambient air 11 is input to a heat exchanger device 71 via inlet24 and hot exhaust gases 91 are circulated back to the heat exchangerdevice 71 under control of valve device 94. The heat from the hotexhaust gases is used to pre-heat the air 29 that is input to thecylinder for combustion. Cooled gases 92 are re-circulated back to theoutput exhaust manifold 35 via exhaust manifold portion 36 for engineoutput. A water recovery unit 66 as described herein is provided tocapture liquid water from any water vapor present in the exhaust gas 91which water may input via a fluid conduit or coupling 68 to the waterreservoir 67.

In the engine 450 of FIG. 7, a dual liquid water and fuel injector 59 isimplemented for directly injecting both fuel and liquid water into thecylinder head portion for combustion. That is, instead of feedingseparate respective liquid water and fuel injectors, liquid waterinjection system 65 and a fuel injection system 55 feed the combinedliquid water and fuel injector 59. Otherwise, the embodiment of FIG. 7is similar to engine 50 depicted in FIG. 3. For example, engine 450 inFIG. 7 also provides for recapturing the heated gaseous product ofcombustion that can be used to further heat the air or fuel or liquidwater at the cylinder. For example, the turbo-compressed air mixture 19at the output of the turbocharger 75 is subject to heat regulation,e.g., heat removal via heat exchanging sub-system 70 that includes aheat exchange device 71 for capturing a substantial portion of the wasteheat produced by combustion and converting it into useful energy tocontrol pre-heating of the air/fuel and water. In one aspect, the heatedexhaust gas 91 is re-circulated from the exhaust manifold 35 viamanifold portion 36 for input to the turbo charger element 75, and inputto heat exchanger 71 for use in pre-heating the liquid water to beinjected and/or for use in pre-heating the air/fuel to be injected. Asshown in FIG. 7, heat energy from the re-circulated exhaust air 91 iscontrollably added under valve 74 control, e.g., via conduit 73 to theheat exchanger 71 to thereby regulate temperature of intake air 29provided to the cylinder head portion 20 for combustion. Further heatedgas exhaust is removed via heat exchanger 71, e.g., via conduit 72. Thecontrollably removed heat from the exhaust gases 91 may be used topre-heat the liquid water in water reservoir 67. Further, in theembodiment of the engine 450 of FIG. 7, a water recovery unit 66 asdescribed herein and the like is provided that will capture water fromany water vapor product from the exhaust gas 91 which water may inputvia a fluid conduit or coupling 68 to the water reservoir 67.

Referring to matrix 500 of FIG. 2, further embodiments of the combustionengines depicted in FIGS. 3-7 contemplate direct liquid water injectionat both the cylinder head portion 20 and at port 38 of the intakemanifold 25 at one or more timed instances during the compression stroketo effectively reduce air temperature and increase density and hence airmass flow rate and power.

FIGS. 1A and 1B depict a direct liquid water injection operation 10 in acylinder with a reciprocating piston shown near TDC employed in theengines 50, 150, 450, for example, of respective FIGS. 3, 4 and 7. Asdepicted in FIG. 1A, the cylinder 12 is shown with a reciprocatingpiston 15 at or near TDC in cylinder 12. In communication with cylinderhead portion 20 is an intake valve 21 timely actuated at each cycle toprovide air from an intake manifold 25 for combustion within cylinderhead portion 20. Likewise, in communication with the cylinder headportion 20 is an exhaust valve 31 timely actuated at each cycle toenable exhaust gas products from combustion (carbon dioxide, air or anyother emissions) to exit the cylinder to an exhaust manifold 35, whereit may exit the vehicle as exhaust or, is captured to perform furtherwork for the engine, e.g., heat up air, fuel, and liquid water. Furthershown in FIG. 1A are embodiments of streams of water 45 exiting directwater injector 46, showing the streams impacting interior surfaces ofthe cylinder, which may include parts of the engine head, valves,cylinder walls, or the piston face. In other embodiments, the wateroutflow from the water injector may be directed in other specificdirections or may be a fine atomized spray that will have minimal impactwith interior surfaces of the cylinder or cylinder head.

The engines of the matrix 500 of FIG. 2 and shown in particularembodiments of FIGS. 3-7 operate at higher compression ratios thanconventional engines. In one embodiment, the “effective” compressionratio of the engines is greater than 13:1, and may be as high as 40:1without the use of any turbocharging, or like techniques that seek toboost compression. Thus, for example, the effective compression ratio ofthe engines described herein are determinable based on an inlet airpressure of being about 1 atm., or less, without use of additionalcompression such as provided by turbocharging.

The engines of the matrix 500 of FIG. 2 and shown in particularembodiments of FIGS. 3-7 employ liquid water injection prior toignition. The injected liquid water cools the air charge during thecompression stroke, reducing compression work and absorbing heat thatwould otherwise be lost to the environment. The effect of the liquidwater added during compression and lean fuel mixtures permits the engineoperation at much higher than conventional compression ratios withoutknock.

Further, in the engines of the matrix 500 of FIG. 2 and shown inparticular embodiments of FIGS. 3-7, one or more of the fuel, air, orliquid water may be controllably heated prior to injection, e.g., byemploying a heat exchanger with the exhaust. The liquid water may beheated, which affects the vapor equilibrium of the liquid water, thedegree of cooling effected by the liquid water, and the rapidity ofsteam formation. In one embodiment, the liquid water may be heated to ata temperature of 25° C. or heated up to a temperature of about 80° C.prior to injection. Alternatively, the liquid water is heated such thatthe temperature of the injected water is greater than about 40° C., or atemperature greater that about 50° C., or a temperature greater thatabout 60° C. or, a temperature greater than about 80° C. or higher.Depending upon the pressure of the injected liquid water, it can beheated to even higher temperatures, within a few degrees less than thecorresponding saturation temperature. For example, for pressures of 10,30, or 50 bar the injected liquid water temperature could be about 150°C., 200° C., or 250° C. respectively.

The sum of these features results in engines 50, 150, 250, 350, 450, ofFIGS. 3-7 running at much higher thermodynamic efficiency thanconventional engines. As a result, the radiator will be much smallerthan is required in conventional engines or not even necessary at all,because the features of the inventive engine manage excess heat muchbetter than conventional engines, and minimize unnecessary heat lossesto the environment. With this combination of factors, a transformationalchange occurs in both specific power and fuel economy for a givendisplacement and RPM.

Further, in the engines of the matrix 500 of FIG. 2 and shown inparticular embodiments of FIGS. 3-7, respective fuel injectors 56 arecontrolled to provide fuel at a ratio of air to fuel that is greaterthan stoichiometric, with a water injector that injects liquid water atone or more instances into the cylinder at any time ranging from about180° to about 30° before TDC during compression, and wherein the amountof liquid water injected is greater than amount of water at thesaturation point of water vapor in the ambient air in the cylinder, andwherein the engine has an effective compression ratio greater than 13:1.

In one embodiment, the amount of water to be injected per cycle iscontrolled in relation to inlet pressure, temperature, relative humidityand current engine operating parameters as compression end pressure,load and rpm. The microcontroller described below with respect to FIG.8, may have necessary data/functions stored as look-up tables and cancompute the water quantity in relation to the sensed input parameters.For example, an amount of liquid water (saturation vapor mass) can beapproximately computed by using ideal gas laws and available airsaturation steam table for injection per cycle. For example, in thecontrol system described herein, an amount of liquid to be injected forcurrent cylinder compression cycle may be determined based on a currentvalue of the ambient temperature, e.g., temperature sensed at the intakemanifold or inlet air temperature at the intake valve at the cylinder,and/or from tabulated humidity/climate data, from which the saturationpoint of water vapor at the sensed temperature is determinable. Thecontrol system adds an additional amount of water (by weight or volume)that will be greater than the minimum amount necessary to completelysaturate the air at the sensed inlet air at temperature.

Water Injection

One factor allowing for greater compression in each of the engineembodiments is internal cooling during the compression cycle by theaddition of liquid water in the engine's cylinder(s) during thecompression stroke. Injecting liquid water into the engine cylinderperforms several important functions. The liquid water internally coolsthe interior of the cylinder during compression by absorbing the heatproduced during the compression. This internal cooling has the effect ofreducing the work required for compression, and also has the effect ofallowing greater compression ratios without engine knock.

The amount of liquid water injected into the engine cylinder during acompression cycle is a function of the saturated water vapor capacity ofambient air, or a function of fuel on a weight basis. In an embodiment,the amount of liquid water injected per cycle is greater than the amountthat is required to saturate air at 20° C. Alternatively, the amount ofliquid water injected per cycle may be about 1.05 to about 10 times theamount of water vapor carried in ambient air in the engine inlet.Alternatively, the amount of liquid water may range from about 20% toabout 800% w/w of the amount of fuel injected per compression stroke.The amount of liquid water injected may be metered to optimize internalcooling, minimization of work required during compression, minimizationof engine knock, and to provide or prevent liquid water droplets mixedwith gases at the moment of ignition. The control of the liquid waterinjection may be based on pressure and temperature sensors in the enginethat are coupled to a computer control system that controls the waterinjection system as described below with respect to FIGS. 8 and 9.

A direct liquid water injector 46 may inject liquid water into thecylinder as a liquid droplet stream, either continuously,intermittently, or as a pulsed stream. In another embodiment, the liquidwater may be injected as a “coarse” spray or an atomized flow. Inanother embodiment, the liquid water may be injected as a streamdirected to cool internal surfaces of the cylinder head and piston top.In yet another embodiment, the liquid water is port injected as anatomized spray into the intake manifold. Any combination of theseembodiments may be used and other embodiments of injecting liquid waterare also possible and within the scope of this concept. In one aspect,the size of liquid droplets will determine rate of heat transfer as theinterface area will increase with decrease in droplet size. The size ofthe droplets is controlled by the liquid water injector settings (andother factors such as timing and metering) and will be controlledaccording to the required rate of heat absorption (evaporation) asdictated by compression temperature rise (sensed through pressure). Inone embodiment, the smaller the required rate of heat absorption, thecoarser (less atomized) could be the injected water spray. In oneembodiment, the size of the average liquid water droplets arespherical-like in a “fine” size ranges from about 0.5 μm to about 25 μmin diameter while the size of the average liquid water droplets in a“coarse” size range may be from about 25 μm to about 100 μm.

In several of the engine embodiments, at least a portion of the liquidwater is injected into the cylinder (by a water injection direct in thecylinder or in the air intake) during the first part of the compressionstroke, between about bottom dead center (BDC) (i.e., about 180° beforeTDC) and about 30° before TDC. The presence of liquid water during thecompression stroke will reduce the work required in the compressionstroke, as the liquid water will absorb latent heat from hot air duringcompression and thus reduce its temperature. The enthalpy ofvaporization, (Δvap) of water at standard pressure is about 40.7 kJ/mol,equivalent to about 2250 J/g. This is the energy required to convertliquid water to steam. The enthalpy of vaporization of liquid waterdepends on pressure and reduces to zero at the critical point, 374.4° C.at 22242 kPa (about 222 bar). For pressures of 10, 30, 50, or 100 bar,the enthalpy of vaporization is about 2015, 1796, 1640, or 1317 J/grespectively, which is still substantial. As pressure is related totemperature, the pressure will be proportionally reduced in spite ofsmall increases in mass and the gas constant R. The reduction inpressure during compression will therefore reduce the work necessary toachieve compression. The work required for compression depends only onpressure as a given volume is fixed (δW=P·dV), where W is the work, P isPressure, and dV is volume differential.

Another effect of added liquid water during the compression stroke is toreduce or eliminate the need for external jacket cooling. The need forexternal cooling arises due to very high gas temperatures and inevitableheat transfer to internal surfaces in the cylinder, particularly thecylinder head, but also the piston face. The injected liquid water mayform droplets in the cylinder that cools the gases in the cylinderduring compression.

Alternatively, liquid water injection may be configured to spray andcool internal surfaces of the engine (cylinder head and piston head)rather than directly cool the gases in the cylinder. In one embodiment,the fuel is expected to be injected predominantly in radial directionand water predominantly in axial direction such that these may notinterfere substantially. An embodiment of this concept is shown in FIG.1A illustrating the centrally mounted liquid water injector 46 oncylinder head 20. The injector is timely actuated under programmed logicor microprocessor control to direct a water droplet stream(s) (e.g., acontinuous stream, an intermittent or atomized spray) of liquid water inthe cylinder head. Representative streams of water are shown as dashedlines 45 in FIG. 1A. As further shown in FIG. 1B, depicting a cylinderin an engine having an intake valve 21 and an exhaust valve 31, such aliquid water injector 46 may be mounted offset from center. Also shownin FIG. 1B is sparkplug 47 and fuel injector 56. In the embodiment ofFIG. 1A, the injector may be actuated to provide a metered volume ofliquid water spray comprising a pattern of one or more individual liquidwater droplet streams 45, aimed at one or more directions andindividually timed to inject water within cylinder head portion 20, orat the piston head at various times during compression.

In operation, some of the liquid water injected may remain unevaporatedas liquid droplets mixed with other gases in the cylinder at the momentof ignition. This liquid water converts to steam after the ignition.Thus, there is provided a higher density medium before the ignition andsubstantially greater expansion of the water droplets leading to higherefficiency. This embodiment is based on the principle that liquid waterexpands in volume about 1600-fold when converted to steam (at 100° C. atstandard pressure). Furthermore, the vaporization of liquid water duringcombustion will provide a denser medium and enhanced expansion pressure,from the greater volumetric expansion of water to steam as compared toother gases in the cylinder during combustion. This embodiment may alsoallow for cooler exhaust gases due to the high latent heat ofvaporization (enthalpy of vaporization) water. In this case, less heatwould need to be rejected to keep the engine temperature within itsoperating limits.

In an alternative embodiment, whether directly input to the port orcylinder, the amount of liquid water added during the compression strokeis metered to minimize the presence of liquid water when combustion isinitiated. This addresses the concern that liquid water present at thebeginning of ignition will absorb heat from the combustion gases,reducing both temperature and pressure in the power stroke, for the samereasons that the work required during compression will be reduced whenliquid water is present—pressure and temperature are reduced from thelatent heat of vaporization as liquid water is vaporized.

In another embodiment, whether directly input to the port or cylinder,the metered amount of liquid water may be injected into the cylinder ata time ranging from about 180 to about 30 degrees before TDC. In theembodiment shown in FIG. 1A, the liquid is intentionally directed to thepiston face and cylinder head and intake and exhaust valve heads, toavoid having the water well mixed with the fuel charge within thecylinder. This embodiment is illustrated in FIG. 1A showing severalcoarse liquid water streams of the spray pattern 45 that do not mix wellwith the fuel charge (fuel/air mixture). This embodiment addresses theconcern that any liquid water present during combustion will reducepressure and temperature during the power stroke. This embodiment solvesthe problem of using liquid water injected strategically to cool theengine, while at the same time minimizing the likelihood the liquidwater will cool the combustion process and reduce the power output ofthe engine. During compression the addition of liquid water to absorbheat from hot air would be optimal if the liquid water was injected intothe cylinder and thoroughly mixed with the gases in the cylinder. Butduring the late part of compression stroke and early part of the powerstroke, near TDC, the cylinder volume and cylinder wall area is verysmall as compared to the combined areas of the piston face and cylinderhead, including valves. By injecting liquid water during late part ofcompression stroke that is directed to the cylinder head or pistonsurfaces, and avoiding mixing of liquid water with hot gases,substantial cooling of the engine can be achieved without cooling thecombustion process post-ignition. This method may be able to capturealmost all the heat internally that was going to the coolant and theradiator. Furthermore, it is well known that the cylinder head is a veryhot portion of any internal combustion engine. The heat absorbed byliquid water internally in this embodiment can be regarded as stored inbuffer to be recovered in the exhaust to heat injected water, fuel andif required, inspired air.

In a further embodiment, the engines employ a mixture of liquid waterwith an alcohol, or other additives that are commonly added to water inengines to lower the freezing point of liquid water, such as, e.g.,methanol, ethanol, isopropanol. The use of such an alcohol additiveprevents the liquid from freezing in cold weather, which is an importantconsideration in for example, automotive applications in cold climates.The liquid water-alcohol mixture ratio may range from about 0% to about50% by weight. As used herein, with respect to the water-alcohol mixtureratio, it is understood that ranging from about 0% means that an amountof alcohol (or like additive) is present.

Lean Fuel Mixture

In combination with various embodiments of the engines as describedherein is the use of lean air/fuel mixtures. That is, the fuel isinjected either into the inlet air stream or injected directly into thecylinder with a fuel injector. The amount of fuel is adjusted to keepthe air/fuel mixture lean. This means that a molar excess of oxygen, inair, is employed in the engines. Stoichiometric air to gasoline fuel isapproximately 14.7:1 (w/w). The ratio of actual amount of air tostoichiometric air for the injected fuel is expressed as λ (i.e., therelative air fuel ratio), where λ=1 is defined as stoichiometric. Asdefined herein, λ>1 is a lean ratio, and λ<1 is a rich (oxygendeficient) ratio.

Combustion efficiency may increase with lean mixtures, because thelimiting reagent is air rather than fuel. At rich mixtures, there willbe non-combusted fuel in the exhaust, which is wasted energy. Combustiontemperatures are also lower with lean mixtures, leading to reduced heatlosses. Of course, there is an optimal air-fuel ratio that depends onthe fuel, temperature, and pressure at the moment of ignition. A keyfeature of the engines described herein is that λ can be increased farmore than in conventional spark ignition engines because the instantengine is capable of operating at much higher compression ratios thanconventional engines.

In one aspect, liquid water injection in the amounts and timing relativeto TDC as described herein achieves more uniform mixing of fuel and airthan in conventional engines. In another aspect, the high compression,along with excess oxygen (air) allow for higher temperature and pressureprior to ignition leading to higher rate and extent of combustion andconsequent higher efficiency. Thus, lean mixtures are expected todecrease engine knock by increasing the auto ignition temperature.Conventionally, gasoline engines will not operate reliably at λ>1.5, butthe engines described herein are expected to operate efficiently atλ>1.5 to as much as about λ=8. In an embodiment, the engines describedherein operate with λ greater than about 1.2; in another embodiment, λis greater than about 1.5. In another embodiment, λ is greater thanabout 2.0. In another embodiment, λ is greater than about 4.0. Inanother embodiment, λ is greater than about 6.0. In an embodiment, theair to fuel ratio is in a range of about λ≧1.2 to about λ≦8.0; or aboutλ≧2.5 to about λ≦5.5; or about λ≧3.5 to about λ≦5.0.

The inlet air or fuel or air/fuel mixture may be heated individually ortogether prior to injection or inspiration into the cylinder. Heatingthe air or fuel or air fuel mixture can transfer useful energy from theexhaust back to the engine. Furthermore, heating the air/fuel mixtureprovides better and more efficient mixing of air and fuel in thecylinder, and heated liquid fuels at or below the critical state areexpected to evaporate more efficiently and mix better with air.Supercritical liquid fuels will flash into vapor and mix very readilywith air. In such an embodiment, a liquid fuel may be subject to atemperature and pressure condition above its critical point, wheredistinct liquid and gas phases do not exist. Supercritical fluids havingproperties between those of a gas and a liquid for use with combustionand/or compression stroke engines include, but are not limited to:Methane, Ethane, Propane, Ethylene, Propylene, Methanol, Ethanol andAcetone. The air or fuel (or both) may therefore be heated to atemperature selected from about 30° C. to about 150° C. In a furtherembodiment, the fuel/air mixture at the air to fuel ratio is heatedbefore the injection of water to a value ranging from about 30° C. toabout 80° C. or ranging from about 40° C. to about 80° C., or rangingfrom about 50° C. to about 80° C., or to a value of about 80° C. orgreater before the injection of water.

Engine Temperature Control

If an embodiment where air or fuel is heated, the heat is supplied froma heat exchanger 71 or like device that captures heat from the exhaustand transfers some of the exhaust heat to the air or fuel. This is anaspect that transfers heat that would otherwise be wasted and lost tothe environment to useful work. Inlet air may also be preheated in coldstart conditions.

In order to control the amount of cooling affected by added liquid waterduring the compression stroke, the engines may employ one or moretemperature and pressure sensors at various locations. As shown in FIG.8 depicting a computer control system 100 for controlling and monitoringengine system operations, temperature sensor devices 110 may be located,for example, in the inlet manifold/cylinder head, a second temperaturesensor device 111 may be located in the exhaust manifold, and a pressuresensor device 112 may be located in the cylinder/inlet manifold/exhaustmanifold, or any combination thereof. Temperature and pressure sensorsmay be located in other locations in an engine as described herein also.Other means of measuring cylinder pressure may be used, including, butnot limited to, combustion chamber plasma monitoring, or crank angleacceleration monitoring. As shown in FIG. 8, one or more temperaturesensors including cylinder head sensor and pressure sensors or sensorsare coupled to a computer employing appropriate software and enginecontrols that can change the air/fuel mixture, the heating of the air orfuel (if used), and amount of liquid water added, depending on factorssuch as the engine temperature or required engine output/rpm. Forexample, an engine that is just started and is running cold may have aslightly richer mixture and less water until warmed up. When fully warm,the air/fuel mixture (and optionally temperature) and added water can bemodulated to adjust the power output and engine efficiency.

The amount of cooling required is based on the maximum operatingtemperature of the various components of the engine, above which somepart will melt or deform or lubrication may fail. The conventionalsolution to engine temperature control is a heat exchanger (radiator)with a fluid (engine coolant) that circulates through the engine andconveys excess heat out of the engine into the environment via theradiator. In practice, the amount of wasted heat in conventional enginesas typically at least 40%. From the Carnot theorem, some of this wastedheat theoretically is thermodynamically available for conversion tomechanical energy.

Thus, in an embodiment, the engines additionally employ sufficientliquid water added during the compression stroke to cool the engine tothe point that a radiator will not be necessary. In a furtherembodiment, the combustion engines described herein provides a method toreduce or eliminate waste heat that needs to be rejected to theenvironment, because of the liquid water injected into the cylinder. Theamount of liquid water would be metered according to data provided bythe temperature and pressure sensors. In an embodiment, the liquid waterinjected into the engine may be preheated by circulation from within theengine, thereby reducing heat losses to the environment. As notedelsewhere, the liquid water may be heated to about 80° C. or even higherin high pressure environments. Further features that are expected toassist in cooler internal temperatures are the lean fuel mixtures(λ>1.5), and also the high effective compression ratios. By the term“internal temperature” is meant the temperature at the cylinder head,which is typically the warmest part of an engine. The high effectivecompression ratios are expected to produce a cooling effect due to thegreater volumetric expansion of the cylinder during the power stroke ascompared to conventional lower compression ratio engines.

If the engines employ a radiator (e.g., a heat exchanger) and liquidcoolant for external cooling, the amount of liquid water injection andother cooling features of the engines are expected to reduce by at least20% the amount of cooling (heat rejected to the coolant) required ascompared to conventional engines. In other embodiments, the amount ofheat rejection by the coolant is reduced by at least 40% by liquid waterinjection as compared to an engine without liquid water injection. Inother embodiments, the amount of heat rejection by the coolant isreduced by at least 60% by liquid water injection as compared to anengine without liquid water injection. In other embodiments, the amountof heat rejection by the coolant is reduced by at least 80% by liquidwater injection as compared to an engine without liquid water injection.In another embodiment, a coolant is employed with higher boiling point,e.g. using higher amounts of glycol or operating the cooling loop athigher pressures to allow for a cycle running at higher temperature.

In an embodiment, the engines may further require no external means ofcooling. In another embodiment, the engines may be air-cooled, lacking aheat exchanger entirely. The amount of air cooling may be controlled byeither controlling speed or intermittency of an electrically powered fanor by other means such as controlling operation of a flap. For example,a flap may control the flow of air, either actively using a flap and afan or passively by controlling the surface areas exposed and flowcontrolled by opening of the inlet of air and exhaust. Another way tocool the engine and recover some of the waste heat is by circulatinginlet air around the engine.

In alternative embodiments, the engines may be designed to operate at ahigher temperature than conventional engines. Conventional engines aretypically set to run at an internal temperature of about 91° C. (195°F.), but the engines of matrix 500 described herein may be set to run atan internal temperature of 100° C. to 175° C. with suitable changes inlubricant specifications. In one embodiment, the engine operates at acylinder temperature ranging from about 85° C. to about 175° C. (i.e.,the external temperature of the engine walls that the coolant orradiator water would experience when a cooling system is employed). Incombination with the additional heat management features describedherein, the engine requires no external means of cooling, butoptionally, can include an exhaust radiator. For example, the engineoperates at an external temperature ranging from about 85° C. to about100° C. or ranging from about 85° C. to about 120° C., or ranging fromabout 85° C. to about 140° C., or ranging from about 85° C. to about150° C. and the engine requires no external means of cooling.

In alternative embodiments, the engines may be insulated to minimizeenvironmental heat loss, with cooling only coming from the injectedwater, and optionally, from heat captured in the exhaust for heating airand fuel. In this embodiment, the engine will be designed to run at ahigher internal temperature than a normal engine. In one embodiment, thecombustion chamber or cylinder, or a portion of the engine housing thecombustion chamber, or the entire engine, is optionally heat insulatedby a heat insulator 90 known to one of ordinary skill in the art.

In alternative embodiments, the engines recover heat in the exhaust orengine head by a heat exchanger 70 that transfers heat from the exhaustor engine head or both to pre-heating of liquid water, fuel and inletair. If the engine has no other external cooling apparatus, such as aradiator, or if the engine is insulated to minimize environmental heatloss, pre-heating of the fuel and liquid water can be a means totransfer heat that would otherwise be lost to the environment throughthe exhaust to useful mechanical energy.

As a result of the cooling measures that control the temperature in thecylinder during the compression stroke, including liquid water injectionand lean fuel mixtures, greater compression ratios are possible thanwith conventional engines. The engines of the present disclosure have aneffective compression ratio greater than 13:1, but more preferably willhave an effective compression ratio of greater than 15:1, or greaterthan 20:1 or greater than 25:1, or greater than 30:1 and can be as highas 40:1. The higher compression ratios achievable by the engines will bemore efficient than conventional engines in part because of the highercompression ratios available by the apparatus and methods. According tothe Otto cycle (or diesel cycle in the case of compression ignitionengines), higher compression ratios theoretically will result in greaterthermal efficiencies.

The fuel used by the engines may be a low alkane, such as natural gas,methane, ethane, n-propane, or isopropane, or lower alkyl aldehyde orlower alkyl ketone, wherein lower alkyl contains 1-6 carbon atoms (e.g.,acetone), or a mixture thereof. Alternatively, the fuel may be gasoline(petrol) optionally mixed with an alcohol, e.g., ethanol. Otherhydrocarbons may be used as fuels in the engines, such as other C4-C15alkanes or mixtures thereof, or diesel (kerosene) fuels. Gasoline andlow alkane fuels normally will likely require spark ignition. Dieselfuels are compression ignited and the engines may use a fuel mixturebased on a diesel-like fuel, e.g., Diesel, Biodiesel, Kerosene, JP-8,JP-A and other Kerosene type fuels. Both types of fuels and ignitionmethods are compatible with the embodiments described. In an embodiment,the fuel may be a mixture of natural gas and a diesel-like fuel, whereinthe diesel-like fuel causes ignition by compression but the bulk of thecharge is from natural gas.

Ignition in the engines is from a spark plug, from compression ignitionor the combination or another other means such as plasma discharge orlaser. In the case of spark ignition, the timing may be varied dependingupon the fuel, air fuel ratio, and amount of liquid water being injectedor any combination thereof. Ignition is timed to be initiated prior toTDC because the entire fuel charge does not ignite instantaneously. Theprocess of combustion once ignition is initiated takes time, as theflame front formed on ignition moves through the cylinder. For thisreason, ignition, however initiated, is timed (e.g. under computersystem control), to ensure that the maximum pressure from combustionoccurs at or slightly after TDC. In spark ignition engines, “sparkadvance” (timing of the spark) is adjusted to optimize ignition timingto maximize efficiency. The engines described herein operating with leanfuel mixtures and high compression ratios may require less sparkadvance, due to better and more uniform mixing of fuel and a smallercombustion compartment, so that maximum pressure from combustion will beachieved more quickly than in conventional lower compression ratioengines.

Compression ignition engines, which have no spark ignition, willtypically require adjustment of the timing of the fuel being injectedinto the cylinder because of the water cooling in the inventive engine.Thus, with higher λ values and more water injection, which equates tocooler internal temperatures, a computer control system described hereinis operated to inject a diesel-type fuel in a compression ignitionengine earlier in the compression stroke in order to get appropriatelytimed ignition and complete combustion.

FIG. 8 further shows a sensor control system 100 that may be employed inthe engines of FIGS. 2-7. The sensor control system 100 dynamicallycontrols the engine operation by a controller device 105 an EngineControl Unit or ECU (e.g., a microprocessor or programmable logiccontroller or microcontroller) operating under program control that isstored in an associated memory storage device. For engine control,sensor devices are employed including, but not limited to: an MAF or(mass airflow) sensor for measuring a mass of air at the intake; an IAT(intake air temperature) sensor, e.g., in or at the cylinder head and/orthe air inlet; an EGT (exhaust gas temperature) sensor, e.g., in theexhaust manifold, an MAP (Manifold Absolute Pressure) sensor in or atthe inlet/exhaust manifold, or any combination thereof. The sensordevices communicate with and are coupled to controller device 105, whichmay be a computer with a microprocessor that sends out control signalsto control and adjust engine parameters programmably in response tovarious engine temperature and pressure measurements, in addition toother relevant data, such as engine load, or outside air temperature andpressure.

For example, operating engine parameters may be dynamically adjustedaccording to power/speed output requirements of the engine (e.g., engineload) and temperature cooling targets in which the programmedmicroprocessor or programmable logic controller element 105 isresponsive to the power setting (load) 113, e.g. such as indicated byengine RPM, and one or more engine operating conditions. For example,messages or information indicating engine operating conditions arecontinually sensed by sensor devices and communicate real-time valuesthat are input to the programmable logic controller element 105 include,but are not limited to: a first temperature T1 value of the cylinder, asecond temperature T2 of the cylinder at the exhaust manifold, andpiston cylinder pressures P to determine the parameters for operation ofthe engine at a next cycle, e.g., parameters for providing control ofair intake valves and/or fuel injector control and parameters forcontrolling liquid water injector/variable pump control.

As described herein, the amount of water injected is an amount that isgreater than the amount of water that is present at the saturation pointof water vapor in the ambient air in the cylinder. This amount isdeterminable by a skilled artisan. Standard textbooks of Thermodynamicscontain a chapter/section on Humid Air. Saturation vapor mass can beapproximately calculated by using ideal gas laws and Saturation Steamtables.

The amount of water vapor is constrained by the restrictions of partialpressures and temperature. Dew point temperature and relative humidityact as guidelines for the process of water vapor in the water cycle. Thebalance between condensation and evaporation gives the quantity calledvapor partial pressure.

The maximum partial pressure (saturation pressure) of water vapor in airvaries with temperature of the air and water vapor mixture. A variety ofempirical formulas exist for this quantity; the most used referenceformula is the Goff-Gratch equation:

${\log_{10}(p)} = {{{- 7.90298}\left( {\frac{373.16}{T} - 1} \right)} + {5.02808\; \log_{10}\frac{373.16}{T}} - {1.3816 \times 10^{- 7}\left( {10^{11.344{({1 - \frac{T}{373.16}})}} - 1} \right)} + {8.1328 \times 10^{- 3}\left( {10^{{- 3.49149}{({\frac{373.16}{T} - 1})}} - 1} \right)} + {\log_{10}(1013.246)}}$

where T, the temperature of the moist air, is given in units of Kelvin,and “p” the partial pressure of water, is given in units of millibars(hectopascals). Thus, at various temperatures, the partial pressure ofwater when the air is fully saturated can be determined using thisformula. The amount that is injected is greater than that amount “p”calculated by this equation. For example, at 101.33 kPa and 20° C. aircan have a maximum of about 1.5% vapor mass as compared to about 6.8%stoichiometric fuel mass. At 25° C. it is about 2%. In an embodiment,for example, the amount of water injected ranges from about 1.05 toabout 10 times an amount of water vapor carried by air saturated withwater vapor at ambient temperature of about 25° C. Thus, whatever valueof “p” is calculated from the equation such as that given hereinabove,in this embodiment, the amount of water vapor injected is about 1.05 toabout 10 times that value. This amount can then be converted to theamount of liquid water to be added by determining the number of molesthis amount in vapor represents from the ideal gas law equation, PV=nRT,where P is the partial pressure of water to be added, V is the volume ofthe cylinder, T is the temperature in Kelvin, R is the ideal gasconstant and n is the number of moles. Based on the number of moles ofwater calculated, one can calculate the amount of water to be injectedin grams of water, since water has a molecular weight of 18 grams/mole.Since water has a density of about 1 gm/mL, one can then calculate theamount of liquid water to add in milliliters.

As described herein above, the optimal amount of water to be injectedinto the engine is calculated, for example, by the ECU (Engine ControlUnit) via one or more equivalency tables. The equivalency tables containinformation regarding how much water to inject into the engine underdifferent operating conditions. One example of an equivalency table is“Water injection amount VS Intake Air Temperature (IAT) Multiplier”.This is very similar to the “Ignition Timing Retard VS IAT multiplier”table electronically controlled engines have, which retard the sparktiming with increasing intake air temperatures to suppress detonation,except that in the present case, the water injection amount ismultiplied by a small positive number that grows with increasing IATs inorder to account for the increased charge cooling demand of a hotterintake charge taking into consideration various parameters, such as therpm of the engine, the load, the temperature, pressure, fuel, and thelike. Any other factor that has an impact on the propensity of knock tooccur will have a table for it in the ECU that looks at the increase inthe property, and counter it by retarding spark timing, increasingfueling, and increasing/decreasing the amount of water. The primarywater injection equivalency table co-relates the amount of water, fuel,and engine load.

The water injection equivalency table is generated experimentally byrunning injection sweeps (holding the engine at a constant speed andload and varying the amount of water injection from 0 to 100%) atvarious speeds and loads so that the optimum amount of water isidentified under most operating conditions. Data is interpolated inbetween test results to produce a full matrix for the points that sit inbetween actual test points, so when the engine runs through variousloads and speeds the ECU knows exactly how much water it needs to injectin order to keep it running optimally.

More specifically, a method 200 for determining an optimal amount ofwater to inject for each piston per engine cycle is described in FIG. 9.At 210, FIG. 9, there is depicted the controller determining currentengine operating conditions including, e.g., engine RPM, Load. Then, at215, the air flow mass (mass of air flowing into the engine) isdeterminable, e.g., from the MAF, or, an equivalency table (not shown)which co-relates manifold pressure (MAP) and engine RPM to determineairflow. Alternately, air flow mass is determinable from a tablerelating MAF to engine speed and intake air temperature. From thisdetermined air flow mass value, an amount of fuel is calculated at 220given the desired A/F (air/fuel) ratio.

Then, continuing to 225, there is determined via equivalency look-uptables a base water injection amount to inject based on the determinedfuel mass. As described herein, the base water amount injected, asdescribed hereinabove, is an amount of water injected that is greaterthan the amount of water that is present at the saturation point ofwater vapor in the ambient air in the cylinder. An example method tocalculate a water injection amount is described herein above.

Continuing, in FIG. 9 at 230, given the current intake air temperaturevalue sensor reading, there is performed by the controller anequivalency table (not shown) look-up to determine a water chargemultiplier adjustment. The multiplier values are experimentallydetermined and are provided in the equivalency table for real-timeadjusting of the base injection amount (e.g., add or remove an amount ofliquid water) given the current intake air temperature value sensorreading. Likewise, at 235, given the current manifold absolute pressurevalue sensor reading, there is performed by the controller anequivalency table (not shown) look-up to determine a water chargemultiplier adjustment. Multiplier values are experimentally determinedand are provided in the equivalency table for real-time adjusting of thebase injection amount (e.g., add or remove an amount of liquid water)given the current manifold absolute pressure value sensor reading. It isunderstood that addition equivalency table look-ups may be performed toadjust water injection amount on a per cycle basis, based on othersensed parameters, e.g., exhaust gas temperature sensor values.

Continuing to FIG. 9 at 240, there is computed a final adjusted waterinjection amount value by applying each of the multiplier adjustments tothe base water injection amount value obtained at 225. Then, at 245, thecontroller consults a further water equivalency chart (not shown) todetermine a trigger (timing) and a dwell time for the water injectordevice to be opened so that the final adjusted amount of liquid waterfor injection (port or cylinder) is provided for the remaining portionof the cycle.

That is, referring back to FIG. 8, in one embodiment, using thepredetermined information 121 stored in one or more equivalency tables120, the logic controller element 105, will compute the controlparameters 125 to effect the engine output conditions such as the amountof liquid water to be injected. These modifications are effected by thecontroller communicating messages 140 for controlling actuation (e.g.,dwell time) of the fuel injector and communicating messages 140 tocontrol the timing of liquid water injection and the amount (volume)(before TDC) of liquid water injection according to the embodimentdescribed herein. At an engine cycle-by-cycle basis, given the currentsensed conditions values, and in response to the current temperature andpressure readings, and other variables, e.g., environmental conditionssuch as ambient temperature, the controller 105 will coordinate theoperation of the system by sending out control messages 140 formodifying the air and fuel injection amounts and timing, and controlmessages 130 that control the amount of liquid water injection (whetherport or cylinder direct-injected) relative to the timing of the sparkignition (advance) at the cylinder during the compression stroke formaximum efficiency, compression and cooling as described herein.

It is understood, that the monitoring and control of the engineoperations at any particular cycle of operation of the engine may beadjusted based on the operation during the prior cycle (including timeaverage of a few prior cycles) to ensure ignition and water injectionsoccurs at the proper crankshaft angle(s) in a stable manner.

Among other benefits, liquid water in the combustion chamber of aninternal combustion engine reduce the internal temperature, which allowshigher compression engines to operate without knock, thus allowing loweroctane fuel to be used in higher compression and more efficient engines.The lower internal temperatures can also avoid and/or reduce NO_(x)emissions, which increase with increased internal temperature. Inaddition, the engines described herein exhibit decreased amount ofcarbon monoxide relative to that produced by conventional engines.

As used herein, a top dead center (TDC) point of each cylindercorresponds to an orientation of the piston as at a point furthest awayfrom the crankshaft within the cylinder. Measured in degrees, a crankangle (referring to the position of an engine's crankshaft in relationto the piston as it travels within the cylinder) for a piston that is attop dead center (TDC) of its compression stroke is zero crankshaft angledegrees. As used herein, a bottom dead center point (BDC) of eachcylinder corresponds to an orientation of the piston closest to thecrankshaft. A crank angle measured in degrees for a piston that is atbottom dead center (TDC) of its compression stroke is at 180 crankshaftangle degrees.

Unless indicated to the contrary, the temperatures used herein refer to° C.

As referred herein, “ambient” is defined as the conditions oftemperature and pressure outside of the energy storage system, e.g.,about 25 degrees C. and 1 atm.

As used herein, the term “hydrocarbon fuel” refers to a fuel comprisedsubstantially of hydrocarbons (more than 80% hydrocarbons by weight),but may additionally include other additives, such as alcohols, e.g.,ethanol.

As used herein, the plural connotes the singular, and vice versa, thesingular connotes the plural.

The following non-limiting examples are illustrative.

Example 1

FIG. 10 depicts a plot of operating engine efficiency at various engineloads, with water injection and supercharged air injection pressure.Operating efficiency was computed based on air/fuel and injected waterinput as described herein. The test engine was a modified Yanmar L100single cylinder diesel engine, coupled to a 5 KW generator having a borex stroke of 86 mm×75 mm, 435 cc displacement, and an engine compressionratio 19:1. The engine was modified with removable plates blocking airflow to facilitate study of internal cooling. Additional holes weredrilled in various locations for a water injector (e.g., by acommercially available fuel injector adapted to inject water at 80 psi),and pressure and temperature instrumentation. Air was injected with asupercharger at the indicated pressures, e.g., 5 PSI (pounds per squareinch), 10 PSI and 15 PSI. Water injection was port injected at 80 psifor 3 ms at 300° before TDC. Engine speed was approximately 3600 RPM.

The electrical output shown is a direct measure of the efficiency of theengine at the various operating parameters. Efficiency was calculated bydividing measured electrical output by the thermal energy content (heatof combustion) of the fuel injected. Fuel was ULSD ultra low sulfurdiesel. Particularly, FIG. 10 plots electrical efficiency of engine atvarious engine loads and three different supercharged air injectionpressures. Table 1 shows the underlying data plotted in FIG. 10, alongwith the ratio of water percent to fuel (by weight), and the λ (theratio of air to fuel). The most efficient data point in Table 1 and FIG.10, at 15 PSI supercharged air pressure and 5037 watts engine load, was55.3 electrically efficient. The water/fuel ratio was 1.5, meaning thatthe ratio was 150% w/w. The λ at the most efficient data point for thisexperiment was 4.46.

TABLE 1 Data for test engine with water injection and supercharging,plotted in FIG. 10 Engine Load (Watts) Electrical efficiency Water/Fuelratio λ 15 PSI Air Injection 1008 18.2 2.4 7.57 2055 32.0 2.1 6.54 301638.7 1.7 5.22 3920 44.4 1.6 4.51 3938 43.9 2.1 4.46 5037 55.3 1.5 4.465020 54.7 2.1 4.38 10 PSI Air Injection 1031 15.0 2.0 5.72 2021 25.8 1.84.74 3045 38.3 1.7 4.89 3976 46.1 1.6 4.32 4785 47.4 1.4 3.65 5 PSI AirInjection 1006 11.2 1.5 4.35 2021 19.0 1.3 3.65 3065 24.3 1.1 3.05 306524.5 1.5 3.05 4138 27.0 0.9 2.45 4145 26.7 1.2 2.43 4956 27.4 0.8 2.02

FIG. 10 shows that at 10 and 15 PSI air pressure with internal cooling,the efficiency increases steadily to 4 KW engine load. At 15 PSI, theoutput increases further to a maximum of 55% efficiency.

Example 2

FIG. 11 and Table 2 show the electrical efficiency of engine at variousengine loads and 10 PSI supercharged air injection pressures, comparingwater injection internal cooling vs. no water injection and air cooling.Water injection was at 80 psi for 3 ms at 300° before TDC. Engine speedwas 6000 RPM. The data in Table 2 shows that at 10 psi air injection,the use of water injection internal cooling dramatically increases theefficiency. At 4 KW engine load, the efficiency increases from 32 to46%. Table 2 shows the underlying data plotted in FIG. 11, along withthe water/fuel ratio, and the λ (the ratio of air to fuel).

TABLE 2 Data for test engine with at various loads and constant airpressure, without and without water injection (FIG. 11) Engine Load(Watts) Electrical efficiency Water/Fuel ratio λ 10 PSI Water Injection3045 38.3 1.7 4.89 3976 46.1 1.6 4.32 4785 47.4 1.4 3.65 No WaterInjection 3044 32.8 0 6.09 4050 32.2 0 4.97 5000 31.5 0 3.84

As the engines described herein may be embodied in different formswithout departing from the spirit or essential characteristics thereof,it should also be understood that the above-described embodiments arenot limited by any one of the details of the foregoing description,unless otherwise specified, but rather should be construed broadlywithin the spirit and scope, as defined in the appended claims.Therefore, all changes and modifications that fall within the metes andbounds of the claims or equivalence of such metes and bounds aretherefore intended to be embraced by the appended claims.

What is claimed is:
 1. An internal combustion engine for use with ahydrocarbon fuel, the engine having at least one cylinder and areciprocating piston therein, an intake manifold in fluid communicationwith at least one air intake valve, at least one exhaust valve in fluidcommunication with an exhaust manifold, and a fuel handling system withat least one fuel injector, a water injector coupled to a water sourcefor injecting liquid water into the port of the intake manifold; aprogrammed control device configured for controlling internal combustionengine operations, said programmed control device receiving one or morereal-time temperature values in said engine and responsively adjustingan amount of water injected and an air to fuel ratio provided to thesaid at least one cylinder, said programmed control device controllingactuation of said liquid water injector to inject a predeterminedquantity of liquid water into the a port of the intake manifold at atime from about 300° to about 180° before Top Dead Center (TDC) of saidpiston during a compression stroke, said liquid water injector injectingan amount of water greater than an amount of water that is present at asaturation point of water vapor in the ambient air in the at least onecylinder, wherein a ratio of air to fuel provided to the at least onecylinder is greater than stoichiometric, and the engine has an effectivecompression ratio greater than about 13:1.
 2. The engine of claim 1,wherein the fuel or the predetermined quantity water or both aredirectly injected into the intake manifold in fluid communication withthe at least one air intake valve.
 3. The engine of claim 1, wherein thefuel is port injected into an intake manifold in fluid communicationwith the at least one air intake valve.
 4. The engine of claim 1 wherethe fuel is directly injected into the at least one cylinder.
 5. Theengine of claim 1, wherein an amount of water injected in an enginecycle is about 1.05 to about 10 times an amount of water vapor carriedby air saturated with water vapor at ambient temperature of about 25° C.at an engine intake.
 6. The engine of claim 1, wherein an amount ofwater injected in an engine cycle is about 20% w/w to about 800% w/w ofthe amount of fuel being injected in the engine cylinder.
 7. The engineof claim 1, wherein the engine is a compression ignition engine.
 8. Theengine of claim 1, wherein the effective compression ratio is greaterthan about 15:1.
 9. The engine of claim 1, wherein the effectivecompression ratio is greater than about 20:1.
 10. The engine of claim 1,further comprising: a temperature sensor in the cylinder coupled to thecontrol device for sensing a real-time temperature value in saidcylinder, a temperature sensor in the exhaust manifold coupled to thecontrol device for sensing a real-time temperature value of combustionexhaust products, and said programmed control device receiving one ormore real-time temperature values from said temperature sensors for saidadjusting.
 11. The engine of claim 1, wherein the engine furthercomprises another water injector coupled to the water source fordirectly injecting liquid water into the at least one cylinder, whereinat least a portion of the directly injected liquid water is directinjected in the at least one cylinder from about 180° to about 30°before TDC.
 12. The engine of claim 1 wherein the predetermined quantityof liquid water is port injected into the intake manifold as an atomizedspray into the port of the intake manifold, said atomized spray beingfinely atomized or coarsely atomized.
 13. The engine of claim 1, whereinthe liquid water is heated prior to injection directly into the at leastone port of the intake manifold, such that the temperature of theinjected water ranges from about 40° C. to about 80° C.
 14. The engineof claim 1, further comprising a radiator containing a fluid coolant,wherein the predetermined quantity of liquid water injected into theport of the intake manifold is sufficient so that heat generated fromthe engine rejected to the coolant is reduced by at least 20% ascompared to the engine run without water injection.
 15. The engine ofclaim 1, further comprising a heat exchanger for transferring heat fromthe exhaust manifold or the at least one cylinder or both to providepre-heating of water, fuel and inlet air.
 16. The engine of claim 1,wherein a fuel/air mixture is modulated in response to the sensedtemperature values and power output requirement of the engine.
 17. Theengine of claim 1, wherein a fuel/air mixture at said air to fuel ratiois created by mixing the fuel and air prior to injection into the atleast one cylinder.
 18. The engine of claim 1, further comprising aturbocharger or supercharger that adjustably modulates a quantity of airforced into an intake manifold or the cylinder.
 19. The engine of claim1, wherein the air to fuel ratio greater than stoichiometric isexpressed as a value λ, said λ being about 1.2 or greater.
 20. Theengine of claim 19, wherein κ is greater than about 1.5.
 21. The engineof claim 19, wherein λ is adjustable within the range of about 1.2 toabout 8.0.
 22. The engine of claim 1, wherein a fuel/air mixture isheated to a value ranging from about 30° C. to about 80° C. before theinjection of water.
 23. The engine of claim 1, wherein the fuel isnatural gas, methane, ethane, n-propane, or isopropane, or a mixturethereof, or said fuel is selected from gasoline, or a gasoline andalcohol blend.
 24. A method of operating an internal combustion engine,said engine using a hydrocarbon fuel with at least one cylinder and areciprocating piston therein, an intake manifold with at least one airintake valve, at least one exhaust valve in fluid communication with anexhaust manifold, and a fuel handling system with at least one fuelinjector, said method comprising: injecting, via a water injectorcoupled to a water source, a predetermined quantity of liquid water intothe cylinder at any time ranging from about 300° to about 180° beforeTDC of said piston during a compression stroke, wherein the amount ofwater injected is greater than the amount of water that is present atthe saturation point of water vapor in the ambient air in the cylinder;adjustably modulating, via a turbocharger or supercharger, a quantity ofair forced into the intake manifold or cylinder during said compressionstroke; wherein a ratio of air to fuel provided to said at least onecylinder is greater than stoichiometric, and the engine has an effectivecompression ratio greater than about 13:1.
 25. A method of operating aninternal combustion engine, said engine using a hydrocarbon fuel with atleast one cylinder and a reciprocating piston therein, an intakemanifold with at least one air intake valve, at least one exhaust valvein fluid communication with an exhaust manifold, and a fuel handlingsystem with at least one fuel injector, said method comprising:injecting, via a water injector coupled to a water source, apredetermined quantity of liquid water into the cylinder at any timeranging from about 180° to about 30° before TDC of said piston during acompression stroke, wherein the amount of water injected is greater thanthe amount of water that is present at the saturation point of watervapor in the ambient air in the cylinder, adjustably modulating, via aturbocharger or supercharger, a quantity of air forced into the intakemanifold or cylinder during said compression stroke; wherein a ratio ofair to fuel provided to said at least one cylinder is greater thanstoichiometric, and the engine has an effective compression ratiogreater than about 13:1